1. Field of the Invention
The present invention relates to a vehicle, in particular a rail vehicle, having a car body, which is supported on a running gear in the direction of a vehicle height axis by means of a spring device, and a rolling compensation device, which is coupled to the running gear and the car body, wherein the rolling compensation device, in particular, is arranged kinematically in parallel to the spring device. The rolling compensation device counteracts rolling motions of the car body toward the outside of the curve about a rolling axis parallel to the vehicle longitudinal axis during travel in curves, wherein the rolling compensation device, for enhancing tilting comfort, is configured to impose, in a first frequency range under a first transverse deflection of the car body in the direction of a vehicle transverse axis, on the car body a first rolling angle, which corresponds to an actual curvature of a track section currently negotiated. The present invention also concerns a corresponding method for setting the rolling angle on a car body of a vehicle.
2. Description of Related Art
On rail vehicles—but also on other vehicles—the car body is generally supported on the wheel units, for example wheel pairs and wheelsets, via one or more spring stages. The centrifugal acceleration generated transversely to the direction of motion and thus to the vehicle longitudinal axis means that as a result of the comparatively high position of the centre of gravity of the car body the car body has a tendency to roll towards the outside of the curve in relation to the wheel units thus causing a rolling motion about a rolling axis parallel to the vehicle longitudinal axis.
Such rolling motions detract from the travel comfort when they exceed certain limiting values. In addition they also constitute a danger of breaching the permissible gauge profile and, in terms of the tilt stability and thus also the derailment safety, a danger of inadmissible unilateral wheel unloading. In order to prevent this, as a rule, rolling support mechanisms in the form of so-called rolling stabilisers are used. The job of these is to offer a resistance to the rolling motion of the car body in order to reduce the latter, but at the same time not hindering the rising and dipping motion of the car body in relation to the wheel units.
Such rolling stabilisers are known in various hydraulically or purely mechanically operating designs. Often a torsion shaft extending transversely to the vehicle longitudinal axis is used, as known from EP 1 075 407 B1, for example. On this torsion shaft, on either side of the vehicle longitudinal axis, levers secured against rotation are located, extending in the vehicle longitudinal direction. These levers are in turn connected to rods which are arranged kinematically in parallel with the suspension devices of the vehicle. When the springs of the suspension devices of the vehicle deflect, the levers located on the torsion shaft are set in a rotational motion by means of the rods to which they are connected.
If during travel in curves a rolling motion occurs with varying spring deflections of the suspension devices on either side of the vehicle, this results in differing angles of rotation of the levers located on the torsion shaft. The torsion shaft is thus loaded by a torsional moment, which—depending on its torsional stiffness—at a certain torsional angle, it compensates by a counter-moment resulting from its elastic deformation, thus preventing a further rolling motion. On rail vehicles fitted with bogies the rolling support mechanism can also be provided for the secondary suspension stage, i.e. between a running gear frame and the car body. The rolling support mechanism can also be applied in the primary stage, i.e. operating between the wheel units and a running gear frame or—in the absence of secondary suspension—a car body.
Such rolling stabilisers are also used in generic rail vehicles, such as those known from EP 1 190 925 A1. On the rail vehicle known from this document the upper ends of the two rods of the rolling stabilisers (in a plane running perpendicularly to the vehicle longitudinal axis) are displaced towards the centre of the vehicle. As a result of this the car body, in the event of a deflection in the vehicle transverse direction (as is caused, for example, by the centrifugal acceleration during travel in curves) is guided in such a way that a rolling motion of the car body toward the outside of the curve is counteracted and a rolling motion directed toward the inside of the curve is imposed upon it.
This rolling motion in the opposite direction serves, inter alia, to increase the so-called tilting comfort for the passengers in the vehicle. A high tilting comfort is normally understood here to be the fact that, during travel in curves, the passengers experience the lowest possible transverse acceleration in the transverse direction of their reference system, which as a rule is defined by the fixtures of the car body (floor, walls, seats, etc.). As a result of the tilting of the car body towards the inside of the curve caused by the rolling motion the passengers (depending on the degree of tilting) experience at least part of the transverse acceleration actually acting in the earth-fixed reference system merely as increased acceleration in the direction of the vehicle floor, which as a rule is perceived as less annoying or uncomfortable.
The maximum admissible values for the transverse acceleration acting in the reference system of the passengers (and, ultimately, the resultant setpoint values for the tilt angles of the car body) are as a rule specified by the operator of a rail vehicle. National and international standards (such as for example EN 12299) also provide a starting point for this.
Here, with the vehicle from EP 1 190 925 A1, it is possible to create a purely passive system, in which the components of the suspension and of the rolling stabilisers are adapted to each other in such a way that the desired tilting of the car body is achieved solely by the transverse acceleration acting during travel in curves.
For such a passive solution, firstly the rolling axis or the instantaneous centre of rotation of the rolling motion must be comparatively far above the centre of gravity of the car body. Secondly, the suspension in the transverse direction must be designed to be comparatively soft, in order to achieve the desired deflections solely with the acting centrifugal force. Such a transversely soft suspension also has a positive effect on the so-called vibration comfort in the transverse direction, since impacts in the transverse direction can be absorbed and dampened by the soft suspension.
These passive solutions have the disadvantage, however, that because of the transversely soft suspension and the elevated instantaneous centre of rotation in normal operation, but also in unplanned situations (e.g. an unexpected stopping of the vehicle on a curve with a high cant) comparatively high transverse deflections in the transverse direction also result meaning either that the typically specified gauge profile is breached or (in order to avoid this) only comparatively narrow car bodies with reduced transport capacity can be constructed.
The problem of large deflections in order to achieve a certain rolling angle can indeed be mitigated by shifting the rolling axis or the instantaneous centre of rotation. But this allows only even lower rolling angles to be achieved passively. Consequently the system stiffens in the transverse direction so that not only reductions in tilting comfort but also reductions in vibration comfort have to be accepted.
The rolling motion adjusted for the bend of the curve currently being traveled and the current running speed (and consequently also the resultant transverse acceleration) on the vehicle from EP 1 190 925 A1 can also be influenced or set actively by an actuator connected between the car body and the running gear frame. Here, from the current bend of the curve and the current vehicle speed, a setpoint value is calculated for the rolling angle of the car body, which is then used for setting the rolling angle by means of the actuator.
While this variant offers the opportunity of creating more transversely stiff systems with lower transverse deflection, it has the disadvantage that the vibration comfort is impaired by the transverse stiffness introduced by the actuator so that, for example, transverse impacts on the running gear (for example when travelling over switches or imperfections in the track) are transmitted to the car body with less damping.
In order to compensate for at least the disadvantages regarding vibration comfort by transversely stiff suspension, in WO 90/03906 A1 for a passive system it is proposed that, kinematically in series with the rolling compensation device, a comparatively short transverse supplementary suspension stage is introduced. The disadvantage of this solution, however, is that firstly due to the additional components it increases the installation space required, and secondly the problems described above of large transverse deflections or reduced transport capacity are present here again.